Rotary transmission leadthrough for high pressures and high relative speeds

ABSTRACT

A rotary joint for the transfer of pressurized working fluid between a stationary machine part and a rotating machine part, comprising a housing associated with the stationary machine part and a shaft associated with the rotating machine part. The shaft extends into the housing and is tightly surrounded by at least one bush accommodated in the housing. The inner surface of the bush and the circumferential surface of the shaft are provided with sealing faces which slide against each other. The bush comprises at least one radial fluid feed bore which is connected, via an annular channel encircling the shaft, with at least one radial bore in the shaft for conveying pressurized working fluid through the bush to a fluid duct extending axially in the shaft. The fluid feed bore is also connected with the radial bore in the shaft. The bush is radially resiliently movable within the housing and includes at least one radial bore separate from the fluid feed bore extending to a relieving channel encircling the shaft, the relieving channel being adapted to be loaded with pressurized fluid.

The present invention relates to a rotary joint for the transfer ofpressurised working fluid from a stationary to a rotating machine partand/or vice versa, comprising a housing associated with the stationarymachine part, a shaft associated with the rotating machine part, whichshaft extends in the housing and optionally also therethrough, and atleast one bush accommodated in the housing, which bush tightly surroundsthe shaft and is mounted sealedly in the housing in such a way as to beradially resiliently movable, the inner surface of the bush and thecircumferential surface of the shaft being provided with sealing faceswhich slide against each other, and the bush comprising at least oneradial bore which is connected, via an annular channel encircling thecircumferential surface of the shaft and/or the inner surface of thebush, with at least one radial bore in the shaft for conveyingpressurised working fluid through the bore in the bush to a fluid ductwhich extends axially in the shaft and is connected with the radial borein the shaft.

Such a rotary joint is known, for example, from GermanOffenlegungsschrift No. 38 06 931. In the case of the known rotaryjoint, it is a question of so improving a rotary joint with radial gapthat said rotary joint operates absolutely reliably under all operatingconditions, especially at high speeds and simultaneous high pressuresand additionally exhibits the lowest possible degree of leakage. This isachieved by radially displaceable mounting of the bush, wherebymanufacturing tolerances and possible vibrations or deviations from theconcentric position of the two sliding sealing faces may be compensatedwithout excessive friction occurring between the sliding sealing faces.Although the interposition of the bush, which is mounted radiallyresiliently and simultaneously sealedly in the housing, makes itpossible to increase the rotational speeds and/or diameter of the rotaryjoints without excessive leakage losses, in many instances of use suchrotary joints still remain a speed- and/or shaft diameter-restrictingelement. With the known solution, it is impossible to achieve reliable,low-friction operation without seizing and with a simultaneously lowleakage rate in particular when the operating pressure is not constant,but rather varies between high and lower values. For this reason, therehas also in some cases been a move towards the repositioning ofcorresponding rotary joints onto a shaft extension portion or end, forexample, which does not transmit torque and is only slightly loaded andwhich may accordingly be provided with a tapered diameter. However, theproblem then arises of appropriately positioning the ducts extendingaxially in the shaft or of effecting connection with the axiallyextending duct portions in the shaft extension zone. The shaft or therotating machine part thereby becomes substantially more complex.Moreover, from time to time it is necessary to effect transmission ofthe working fluid in a torque-transmitting shaft portion with acorrespondingly large diameter.

In the case of such rotary joints, the so-called sliding sealing facesbetween shaft and bush are also not provided for the purpose of actuallysliding frictionally on each other, but rather there should if possiblebe provided between these ideally concentric, cylindrical sealing facesa very narrow but uniform gap through which working fluid or lubricantpasses and through which there flows as slight as possible a leakageflow. The narrower this gap, however, the more likely it is that directcontact between shaft and bush will arise and, in the worst case,seizing thereof. With larger gap widths, on the other hand, largeleakage losses arise, which are no longer acceptable, especially withhigh pressures and large shaft diameters.

With respect to this prior art, the object of the present invention isto provide a rotary joint with the above-mentioned features, whichpermits a further increase in shaft diameter, in the relative speeds ofthe surfaces sliding against each other and/or in the pressures,especially even varying pressures, at which working fluid may besupplied.

This object is achieved in that a relieving device is provided byconstructing the bush with at least one radial bore separate from theworking fluid feed bore, which radial bore extends to a pressure chamberformed either on the inner surface of the bush or the outer surface ofthe shaft and may be loaded with fluid pressure independently of thesupply of working fluid through the feed bore.

In the sense of the present description, working fluid is a fluidprovided for any desired purpose, e.g. hydraulic oil, a rinsing agent,water or compressed air. The term "working fluid" is used heresubstantially only as a distinction from the relieving fluid, which may,however, be identical to the working fluid. It is also unnecessary forthe shaft to be a torque-transmitting component in the conventionalsense, but rather "shaft" in this description merely designates arotating, cylindrical machine part. Moreover, the roles of the rotatingand stationary machine parts could even be exchanged, such that theshaft would become the stationary machine part.

In such rotary joints the so-called sliding sealing faces are also notprovided for the purpose of actually sliding frictionally on each other,but rather there should if possible be provided between these ideallyconcentric, cylindrical sealing faces a very narrow but uniform gapthrough which working fluid or lubricant passes and through which thereflows as slight as possible a leakage flow. The narrower this gap,however, the more likely it is that direct contact between shaft andbush will arise and, in the worst case, seizing thereof. With larger gapwidths, on the other hand, large leakage losses arise, which are nolonger acceptable, especially with high pressures and large shaftdiameters.

The pressure chamber formed between the inner surface of the bush andthe outer surface of the shaft, which chamber may be loaded withpressure independently of the supply of working fluid into the shaft,ensures a certain expansion of the bush and reduces the contact pressureand friction between the sealing faces sliding on each other even whenworking fluid is not being supplied or is not under high pressure. Inconventional rotary joints, one problem consists, inter alia, in thefact that, although the friction between the sliding sealing faces isstill acceptable during the supply of working fluid and at the same timethe friction heat arising is also partly conveyed away by the workingfluid, the friction and the friction heat arising increase very rapidlyif the supply of working fluid is interrupted or the pressure of theworking fluid is considerably reduced. In many instances, however, thesupply of working fluid is only provided during certain, frequentlyshort periods, the rotating machine part continuing to rotate constantlyand the sliding sealing faces thus rubbing against each other.

Through the independent supply of pressurised fluid into the relievingchamber, which is provided in the form of a recess or groove in one orboth of the sliding sealing faces, not only is a lubricating filmmaintained permanently between the sliding sealing faces but also thepressure prevailing in the relieving chamber itself and between theadjacent sliding sealing faces expands the bush slightly and forces thesliding sealing faces so far apart that they barely touch each other andthe friction and resultant friction heat are considerably reduced. As inconventional rotary joints, the fluid penetrating between the slidingsealing faces is collected axially externally of the bush in leakagecollecting chambers provided accordingly and returned to a leakage tankor the like. The bush may be even more tightly fitting than in knownrotary joints, since the fluid supplied to the relieving chamberprevents or reduces friction between the bush and the rotating machinepart. Thus, leakage losses are reduced to an acceptable level even athigh pressures and with large shaft diameters and/or high relativespeeds between the sliding surfaces. As far as possible it isadvantageous for the same fluid to be used as relieving fluid as servesas working fluid. This is especially true of fluids with lubricatingproperties.

In a first embodiment of the invention, the relieving chamber isconstructed as an annularly encircling groove, either in the outer orcircumferential surface of the shaft or, as is preferred, in the innersurface of the bush. The uniform pressure prevailing in this relievingchamber over the circumference ensures a uniform, if extremely slight,expansion of the bush and also has a certain centering effect betweenbush and shaft, which may possibly be connected with the fact that therotating shaft entrains the fluid out of the relieving chamber into thesealing gap along the sliding sealing faces between shaft and bushand/or that the pressure in the broader gap reduces more rapidly towardsthe leak. The resultant pressure profile has the effect of centering thegap-forming sealing faces on shaft and bush.

In another preferred embodiment of the invention, this relieving grooveis subdivided into several separate segments, i.e. several radial feedbores for the relieving fluid are provided, which are each connectedwith their own relieving chamber or groove extending over a segment ofthe circumference. At least three such relieving chambers are preferablyprovided, the centres of which are each staggered by 120° relative toeach other and which extend over a circumferential segment ofapproximately 100 or 110°, for example, while the remainingcircumferential segments separate the individual chambers from eachother. Similarly, four or more identical chambers could be distributeduniformly and in segments about the circumference of the shaft. Thesegment-type groove portions may optionally be provided as recesses inthe sliding sealing faces in the inner wall of the bush and/or also inthe outer wall of the shaft.

In this connection, an embodiment of the rotary joint which isparticularly effective is one in which the feed ducts for the relievingfluid, preferably the radial feed bores, each contain a throttle, whichis so set or settable that the supply of relieving fluid is very greatlyrestricted by this throttle and an accumulation of pressure with respectto the feed pressure existing upstream of the throttle occurs in thegroove segment supplied with this relieving fluid if the sealing gapbetween the sliding sealing faces and adjoining this groove segment hasa maximum width. If the sealing gap is narrower, the pressure in thecorresponding groove segment is correspondingly greater. In this way,the throttling ensures a very good centering effect for the shaftrevolving in the bush, since, at the point where the shaft lies moretightly against the sliding sealing faces of the bush and wherepotentially greater friction and greater friction heat arise, in thegroove segment arranged on this side there is a higher pressure than inthe other groove segments, which pressure forces the bush away from sideof the shaft. The sealing gap adjoining the other groove segmentsthereby becomes smaller, such that there too a higher pressure buildsup, so that finally, in the case of a rotating shaft, a dynamicequilibrium arises and the shaft is optimally centred and exhibitsminimum friction with the sliding sealing faces of the bush.

The relieving grooves or groove segments are preferably arranged inpairs and provided in the sliding sealing faces of the bush and/or theshaft in the axial direction on both sides of the feed bores or groovesfor working fluid.

Furthermore, decoupling grooves are also provided in the slidingsurfaces, preferably of the bush, between the relieving grooves orgroove segments and the feed bore or feed groove for working fluid,which decoupling grooves are connected with a leakage duct or a leakagedischarge and ensure that the pressurised working fluid does notpenetrate through the sealing gap into the grooves or groove segments ofthe relieving device, thereby impairing or preventing the centeringeffect of the relieving grooves. The decoupling groove ensures thatbetween the grooves or groove segments of the relieving device and thedecoupling groove there is always a pressure difference, i.e. withrespect to the leakage chamber, which is greater, the narrower thesealing gap between the sealing faces in the area between the respectivegroove segment and the relieving groove. Otherwise, the connection,formed by the sealing gap, with the pressure of the working fluid couldimpair the centering effect of the groove segments. Working fluid whichmay escape from the feed bore or feed groove for working fluid in thedirection of the relieving groove is collected by the decoupling grooveand conveyed away to a leakage collecting chamber. Otherwise, theconnection, formed by the sealing gap, with the pressure of the workingfluid could impair the centering effect of the groove segments.

It goes without saying that the decoupling grooves are also preferablyin pairs and separate the grooves for the working fluid from the pairscomprising relieving grooves or groove segments.

The pressure source for the relieving fluid may be fundamentally thesame as the pressure source for the working fluid and accordinglyworking fluid and relieving fluid are also preferably identical. Itcould accordingly be ensured by means of valves that the fluid supply tothe bores and grooves for working fluid may be interrupted, while at thesame time the supply to the relieving grooves or groove segmentscontinues and if need be is stopped when the rotating machine part comesto a standstill.

However, it may be advantageous to provide a separate pressure sourcefor the relieving fluid, especially when the pressure source for theworking fluid is not designed for a constant pressure or is beingchanged or when independent control of the pressure in the relievinggrooves is deliberately desired.

Further advantages, features and possible applications of the presentinvention will become apparent from the following description of apreferred embodiment and the associated FIGS. in which:

FIG. 1 shows a longitudinal section, containing the axis, a completerotary joint according to the present invention;

FIG. 2 shows an axial longitudinal section only through a bush, whichshows more detail than may be seen in FIG. 1; and

FIG. 3 is a cross section through a bush perpendicular to its axis andin a plane which comprises relieving grooves.

FIG. 1 shows a stationary machine part 1 and a rotating machine part inthe form of a shaft 2. The stationary machine part 1 consistsessentially of a housing 11 with radial bores 12 and 14 and an axialbore 13, further radial and axial bores being provided outside thecross-sectional plane illustrated. The housing 11 further accommodatestwo sliding sealing bushes 3, 3', which are arranged with slight radialplay inside the housing and which are kept substantially centred andsealed in the housing by resilient O-ring seals 6, but may beresiliently moved in the radial direction when the O-rings are slightlydeformed. The O-rings are accommodated in corresponding grooves in thecylindrical inner wall of the housing 11, but it is possible also toprovide these grooves in the outer surfaces of the bushes.

The bushes 3, 3' are fixed in their axial position by an inwardlyprojecting shoulder, not described in any more detail, between thebushes and two circlips, likewise not described in any more detail. Twoball bearings 4 are also arranged in the housing 11 axially externallyof the bushes, which ball bearings may primarily absorb radial, but alsoaxial, loads. The closure of the rotary joint is formed in the axialdirection by two sliding seals with resilient sealing lips, which reston the surface of the shaft 2 with only slight force, however, and thusdo not cause any great friction or friction heat.

The radial bores 12, 14 in the housing 11 and also the radial bores 31in the bushes 3, 3' each open on the inside of the housing and theoutside and inside of the bushes in encircling grooves 32 and 33respectively, such that fluid may be conveyed via these bores and thegrooves irrespective of the relative positioning of the radial bores 14,31 in the circumferential direction in the housing or the grooves.Likewise, fluid is conveyed from the bore 31 via the groove 32 into theradial and axial bores 22, 23 in the shaft 2, such that again it doesnot depend on the position thereof in the circumferential direction.

The two bushes 3 and 3' are identical with respect to the supply ofworking fluid and all other, above-described features. However, the bush3 is distinguished from the bush 3' by a relieving device, which isformed by grooves 34 and additional radial bores 5 in the bush 3, viawhich pressurised relieving fluid is supplied. It goes without sayingthat the bush 3' could also be provided with a corresponding relievingdevice, but the combination of a bush 3 with relieving device and a bush3' without relieving device, as shown in the embodiment according toFIG. 1, is also sensible for certain applications, if, for example, theworking fluid (for example water acting as rinsing medium) suppliedthrough the bores 14, 31 and thus through the bush 3' is or has to besupplied under substantially lower pressure than the working fluid (e.g.hydraulic oil) supplied through the bush 3. In this case, the slidingsealing faces between the bush 3', i.e. the cylindrical surface 38thereof, and the circumferential surface of the shaft 2 may be producedwith somewhat greater play in this area, because even with relativelylarge sealing gaps only a slight leakage flow runs off owing to the lowpressure of the fluid supplied, such that in this case the frictionarising and the friction heat are lower. Even if the pressure of theworking fluid is permanently high, it is optionally possible to dispensewith the separate relief, because the fluid penetrating into the sealinggap also has a similar effect to the fluid supplied from a separate,encircling relieving groove. The relieving device is necessary, however,or at least sensible, if the relative speeds of the surfaces slidingagainst each other and the pressure of the working fluid are very high,but can also occasionally exhibit low values, i.e. especially in thecase-of very rapidly rotating shafts or in the case of moderatelyrapidly rotating shafts with large diameters, which have, for example,to be supplied with hydraulic oil or the like which is occasionallyunder very high pressure but is often or sometimes also under very lowpressure. The relieving fluid is conveyed via the axial bore 13 and theradial bore 12 in the housing to the radial bores 5 in the bush 3,wherein here also grooves 35 encircling the inner surface of the housing11 at the orifice of the radial bore 12 again ensure supply of the fluidirrespective of the position of the bore 5 in the circumferentialdirection.

The relieving bores in turn open on the inside of the bush 3 inencircling grooves 34 or in groove segments, each groove segment thenbeing supplied via its own radial feed bore 5. Further details of suchan embodiment are shown in FIGS. 2 and 3. FIG. 2 again shows a bush 3,in the same sectional plane as is the case in FIG. 1 but without housing11 and shaft 2, the sectional plane passing through two diametricallyopposed feed bores 31 for working fluid. This feed bore 31 opens on theinside of the bush 3 into an encircling groove 32, from which fluid maypass into the radial bore 23 in the shaft 2 shown in FIG. 1.

The relieving device is provided in pairs on both sides of the feedbores and grooves for the working fluid, but in FIG. 2 it is onlyprovided with reference numerals on the right-hand side and is describedbelow. In detail, the relieving device comprises three bores 5 which arestaggered relative to each other at circumferential distances of 120°and each open on the inside into a groove segment 34'. In each of thebores 5 there is arranged a throttle 7, which may optionally beadjustable.

The sealing gap 40 between the outer circumferential cylindrical surface28 of the shaft 2 and the inner cylindrical surface 38 of the bush 3 isshown very exaggeratedly in FIG. 3. In fact, the surfaces 28, 38 adjoineach other with a precise fit and a narrow sealing gap, but the pressurein the groove segments 34' ensures a certain expansion of the bush 3 andreduces the frictional engagement between the surfaces 28, 38. This isthe precise purpose of the relieving device, this relieving effect andthe reduction in friction being clearly noticeable even when, instead ofa plurality of groove segments 34', a single encircling groove isprovided for which only a single feed bore 5 need accordingly beprovided. The throttle 7 could then also be dispensed with.

The three groove segments 34' or the relieving fluid accommodated underpressure therein additionally exert a centering action on the shaft 2.This, of course, requires the throttles 7 to be so adjusted that amarked reduction in pressure occurs in the chamber 34' because of thefluid flowing off through the sealing gap 40, if, owing to radialdisplacement of the shaft 2 relative to the bush 3, the sealing gap 40adjoining the relevant groove segment 34' in the axial directionexpands. If, conversely, the shaft 2 is displaced in the direction ofone of the groove segments 34' by exploiting the available radial play,the sealing gap 40 adjacent in this area diminishes in size, so thataccordingly less relieving fluid is able to escape through the sealinggap 40 and the relieving fluid flowing under pressure through thethrottle 7 increases the pressure in this groove segment 34' until anequilibrium is reached between the fluid flowing off through the sealinggap and the fluid flowing through the throttle 7. If the pressure in oneof the groove segments 34' increases in relation to the other two groovesegments 34', the higher pressure in the first groove segment 34'ensures radial displacement of the shaft 2 towards the centre. In thisway, the pressure in the first groove segment 34' is lowered andincreased in the other two segments 34' and an equilibrium is finallyreached, at which the shaft 2 is optimally centred. Although slightfluctuations and vibrations of the shaft 2 about the ideally centredposition may occur, the friction arising between the surfaces 28, 38 isaltogether minimised by this centering effect.

A further feature, not visible in FIG. 1, of the bush 3 is in the formof a pair of decoupling grooves 36, of which, once again, only thatpresent in the right-hand half of FIG. 2 is provided with referencenumerals and described.

The left-hand half of the bush 3 is constructed as a mirror image of theright-hand half. The decoupling groove 36 provided on the inner surfaceof the bush 3 and specifically as a recess in the sealing face 38 isarranged in the axial direction between the feed groove 32 for theworking fluid and the groove segments 34' for the relieving fluid. Thisrelieving groove 36 is connected with bores or ducts 37 which extendradially and axially in the bush 3 and lead to leakage collectingchambers 26' shown in FIG. 1, which lead via bores 26 in the housing 1to a leakage tank 8.

As may be seen in FIG. 2, approximately 2/3 of the sliding sealing face38 of the bush 3 is covered by the relieving fluid, which, starting fromthe groove segments 34', escapes either via the decoupling groove 36 orthe front end of the bush 3 into the leakage collecting chamber 26'. Byway of approximation, it may be assumed that the pressure drops as therelieving fluid travels from the groove segment 34' to the relievinggroove 36 or to the free front end of the groove 3 in approximatelylinear relationship with the distance to the groove segment 34', whilein the area of the groove 34 or 34' itself the full pressure prevails.Thus, in the centre somewhat more than half the pressure prevailing inthe groove segment 34' acts on the surface loaded by the relievingfluid. If working fluid is additionally supplied through the bore 31 andthe feed groove 32, a leakage flow of this working fluid runs off viathe sealing gap between the face 38 and the shaft to the decouplinggroove 36 and also loads this part of the sealing face 38 withcorresponding pressure. This additionally contributes to a reduction infriction between the sliding sealing faces 38, 28, the pressure exertedby the segments 34' and the adjacent portions of the sealing gap 40additionally effecting centering of the shaft 2. With such an embodimentit is possible to operate rotary joints at high speeds or joint radiiand very high transmission pressure, without excessively high losses ofworking fluid arising in the sealing gap 40. The embodiment according tothe invention with separate relieving grooves or groove segments has theeffect that the bushes may be conformed to the circumference of theshaft with relatively tight tolerances, whereby the leakage flow is keptto a minimum even at high transfer pressures, the pressure relief devicenevertheless helping at the same time to prevent the occurrence ofexcessive friction or friction heat. Additional friction-reducingeffects may of course be achieved by conventional measures, such asspecial sliding surface coatings for example. Thus, for example, thecircumferential surface of the shaft could be provided with a ceramiclayer, while the inner surface 38 of the bush could be specially linedwith an anti-friction alloy.

What is claimed is:
 1. A rotary joint for the transfer of pressurizedworking fluid between a stationary machine part (1) and a rotatingmachine part (2), comprising a housing (11) associated with thestationary machine part (1), a shaft (2) associated with the rotatingmachine part, which shaft (2) extends into the housing (11), at leastone bush (3) accommodated in the housing (11), which bush (3) tightlysurrounds the shaft (2), the inner surface of the bush (3) and thecircumferential surface of the shaft (2) being provided with scalingfaces (38, 28) which slide against each other, and the bush (3)comprising at least one radial fluid feed bore (31) which is connected,via an annular channel (32) encircling the shaft (2), with at least oneradial bore (23) in the shaft (2) for conveying pressurized workingfluid through the bush (3) to a fluid duct (22) which extends axially inthe shaft (2) and is connected with the radial bore (23) in the shaft(2), characterised in that the bush (3) is radially resiliently movablewithin the housing (11) and includes at least one radial bore (5)separate from the fluid feed bore (31), which radial bore (5) extends toa relieving channel (34) encircling the shaft (2), the relieving channel(34) being adapted to be loaded with pressurized fluid.
 2. A rotaryjoint according to claim 1, characterised in that the relieving channelis constructed as an annularly encircling groove (34) formed in theinner surface of the bush (3).
 3. A rotary joint according to claim 1,characterised in that the bush comprises a plurality of radial bores(5), which are arranged in various circumferential portions and whichare each connected with a chamber extending over a circumferentialsegment.
 4. A rotary joint according to claim 3, characterised in that athrottle device (7) is provided in the fluid supply for each of thechambers (34').
 5. A rotary joint according to claim 2, characterised inthat a radial bore is provided in the axial direction on both sides ofthe fluid feed bore provided for the working fluid.
 6. A rotary jointaccording to claim 1, characterised in that an annularly encirclingdecoupling groove (36) is provided between the and the fluid feed bore,which decoupling groove (36) is open to the outside or is connected witha leakage collecting chamber (8) via bores (37).
 7. A rotary jointaccording to claim 3, characterised in that the relieving chambers areloaded from the same pressure source as is provided for the workingfluid supply.
 8. A rotary joint according to claim 1, characterised inthat a separate pressure source is provided for the radial bore, whichpressure source preferably supplies the same working fluid as is alsosupplied to the fluid feed bore (31) for working fluid.
 9. A rotaryjoint according to claim 1, characterised in that the relieving channel(34) is formed on the inner surface of the shaft (2).
 10. A rotary jointaccording to claim 1, characterised in that the relieving channel isformed on the outer surface of the bush.
 11. A rotary joint according toclaim 1, characterised in that the annular channel (32) is formed in thecircumferential surface of the shaft (2).
 12. A rotary joint accordingto claim 1, characterised in that the annular channel (32) is formed inthe inner surface of the bush (3).
 13. A rotary joint according to claim1, characterised in that the shaft (2) extends through the housing (11).